EP1775533A2 - Method and apparatus for operating a compression type refrigeration system - Google Patents

Abstract

The method involves determining a control value for a throttle body based on the deviation of actual overheating of a cooling agent from a reference-overheating, and detecting liquefier pressure. Evaporator pressure is measured, and a model comparing the cooling agent mass flow rate at an evaporator input with the mass flow rate at an evaporator output is formed. Another control value is evaluated from the evaporator and liquefier pressures and cooling circuit specific values. Third control value is determined by combining the two values and the body is adjusted to the third value. An independent claim is also included for a device for controlling a compression refrigerator.

Description

The invention relates to a method and a correspondingly designed device for controlling a refrigeration cycle with a refrigerant, an evaporator, a pressure increasing unit, a condenser and a throttle body.

In a compression refrigeration system, the refrigerant present in the refrigeration circuit of the compression refrigeration system is in principle evaporated in the evaporator by heat removal of the medium to be cooled. In the compressor there is a pressure and thus a temperature increase. Subsequently, the refrigerant is liquefied in the condenser with release of heat through the throttle body, the refrigerant is expanded to the evaporation pressure.

Such compression refrigerators are e.g. used for the heating of rooms and the preparation of service water; Both are referred to below as a heat sink.

The regulation of the heat sink temperature is usually carried out by switching on and off of the compressor or by modulation of the compressor speed. Such methods are for example from EP 1 355 207 A1 or DE 43 03 533 A1 known. Furthermore, it is an object of the scheme to optimize the efficiency of the evaporator and thus the refrigeration circuit. The efficiency of the evaporator depends inter alia on its degree of filling, that is, which part of the evaporator with wet steam and which part of the evaporator with superheated Refrigerant medium is filled. The higher the wet steam content, the lower the overheating and the better the efficiency.

However, if the entire evaporator is filled with wet steam and does not get superheated wet steam in the compressor, this can lead to compressor damage. But too small a filling amount of refrigerant in the refrigerant circuit can affect the efficiency of the refrigerant circuit unfavorable, so that an efficiency-optimized degree of filling of the evaporator with wet steam can then no longer be guaranteed.

As a controlled variable for the evaporator control, the overheating of the refrigerant at the evaporator outlet is preferably used. This superheating of the refrigerant can preferably be determined from the evaporator pressure p ₀ and the temperature T _{on of} the superheated refrigerant at the evaporator outlet. Temperature and pressure can be easily measured by suitable sensors. The difference between the evaporator outlet temperature T _oh and the evaporating temperature T ₀ , which is the temperature of the refrigerant during evaporation without overheating, is calculated and is the actual superheat ΔT _{0h-ist of} the refrigerant.

The set value for the evaporation overheating can be set as a fixed value for the refrigeration system. However, it is advantageous to adapt this to the operating point of the refrigeration system. This can be done via a characteristic field or an automatic adaptation as a function of dynamically variable variables in the cooling circuit. For example, when the tendency to oscillate occurs in the control loop, the superheat setpoint can be increased.

An overheating controller then determines the difference between overheating actual and setpoint. Depending on the control deviation, the manipulated variable, here the throttle body, is set.

It has been shown that in practical operation, especially in a wide range of permissible evaporator and condenser temperatures, the refrigeration circuit is exposed to very different working conditions. control engineering seen varies depending on the particular operating point to be controlled route, the refrigerant circuit, strong in gain and offset to adjust the target overheating then also varies the control signal accordingly in a wide range.

If such a refrigeration circuit is regulated, for example, with a conventional controller with preset controller parameters, an exact control is not possible independently of the respective Kättekreisarbeitspunkt because the controller does not adapt to the working point-dependent varying route. Furthermore, in this case it is not possible to output an estimated control signal at compressor start and at this time not yet present overheating-relevant process data.

The invention has for its object to provide a method and a compression refrigeration system of the type mentioned, in which avoided the disadvantages of the above control method and the superheat of the refrigerant optimally controlled at the evaporator output, thus optimizing the efficiency is achieved.

This object is achieved by a method with the following steps: a) determining a first control value for the throttle body as a function of the deviation of an actual overheating of the refrigerant from a target overheating, b) determining the condenser pressure, c) measuring the evaporator pressure d) forming a model comparing the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, e) calculating a second control value for the throttle element based on the model from the evaporator pressure, the condenser pressure and cold cycle specific variables, f) determining a third control value for the throttle element by linking the first control value with the second control value and g) setting of the throttle body to the third control value,

In addition, the object is achieved by a correspondingly designed compression refrigeration system according to claim 14.

The evaporator pressure is a characteristic of the refrigeration cycle size, from which, as well as from the condenser pressure, conclusions about the state of the refrigerant circuit can draw on the basis of basic equations describing the refrigerant mass flow at the evaporator outlet and evaporator inlet, a model is developed according to the invention, which second control value for the throttle body generated. If the first control value, which is determined from direct measured variables of the circuit, linked to the second control value, there is a third control value for controlling the throttle body, which optimally controls the throttle body.

The invention is thus based on the assumption that the functions of the components located in the refrigeration circuit evaporator, compressor, condenser and throttle body can be described approximately using simplified physical Beschreibungsformein.

From a few easily measurable process values can then be calculated on the basis of the model more difficult to determine process variables, in particular the second control value of the throttle body. If this second value according to the invention flows into the calculation according to the invention of the third control value according to the invention by the superheating controller, the precalculated value advantageously results as a well-approximated starting value for the control signal of the throttle element at the start of the compressor.

In a particularly preferred embodiment, the method according to the invention can comprise an expansion valve, a piston engine or a turbine as throttle element.

Instead of the process variable of the evaporator pressure, the process variable of the evaporator inlet temperature can also be used for modeling if it is converted into the evaporator pressure by means of the refrigerant characteristic curve.

In the event of rapid disturbances in the system (rapid changes in the operating point of the refrigeration circuit, for example due to temperature jumps), the second actuating signal reacts immediately. Due to the precalculation of the manipulated variable, the control loop gain is defined and the controller can be adapted accordingly.

The advantages of this second control value according to the invention are that it reacts quickly to changes in the ambient conditions, it is a good starting point for the start of the compression refrigeration system and serves as a reference for a refrigerant deficiency detection,

Based on the variables determined in the method, it is possible to determine a refrigerant shortage. This is determined if, during normal operation, the first manipulated variable exceeds a limit value for a parameterized period of time. Corresponding measures can then be initiated immediately in order to restore the optimal operation of the compression refrigeration system as quickly as possible.

When starting the system, during the time window after and during emergency operation, the throttle can be set to the second control value. Immediately at the start there is still no suitable first control value - derived from the control deviation of the superheat - so the third control value is formed exclusively from the second control value.

An offset of the throttling element, a cold circle specific constant and an exponent are included in the modeling as cold circle specific variables. They are predetermined and characteristic for a cycle, which makes integration in the model easy, since they are entered only once.

In a preferred embodiment, the first control value and the second control value are linked by multiplication. The multiplicative link leads to a simplification of the operating point-dependent evaluation of the refrigerant shortage detection. Furthermore, the multiplicative linkage of the operating point-dependent contributes Line gain calculation and results in an approximately constant gain in the entire control loop.

For special modes, such as defrost or standby, the throttle can be set to a fixed value. An adjustment of the throttle body to predetermined values in the special modes is useful in terms of refrigeration, in order to ensure efficient operation and to condition the refrigerant circuit for the resumption of normal operation.

In a further preferred embodiment, the condenser temperature of the compression refrigeration system is measured and from this the condenser pressure is calculated. The further process steps are identical to the steps a) and c) to g).

In a preferred embodiment, the condenser pressure is measured.

The cold-circle-specific constant enters the modeling as a characteristic variable. It can be determined in laboratory tests for the respective plant or type of plant or, preferably, adapted during normal operation.

The process steps are always performed when the refrigeration cycle is controlled for optimal overheating. This is preferably done regularly, in particular continuously, during the operation of the compression refrigeration system.

In a particularly preferred embodiment, a heat pump is used as a compression refrigeration system

Fig. 1

eine schematische Darstellung einer Kompressionskälteanlage gemäß einem ersten Ausführungsbeispiel;

Fig. 2

eine Darstellung des Ablaufschemas des erfindungsgemäßen Verfahrens;

Fig. 3

eine weitere Darstellung des Ablaufschemas des erfindungsgemäßen Verfahrens;

Fig. 4

eine schematische Darstellung einer Kältemaschine gemäß einem zweiten Ausführungsbeispiel;

Fig. 5

eine Darstellung des Ablaufschemas zur Berechnung einer gemittelten Tautemperatur;

Fig. 6

eine Darstellung des Ablaufschemas in Abhängigkeit von der gemittelten Tautemperatur und der Zeit;

Fig. 7

eine Darstellung der berechneten Werte für die Amplitude in Abhängigkeit von der Zeit;

Fig. 8

eine Darstellung der Außentemperaturkompensation;

Fig. 9

eine Darstellung des Temperaturverhaltens von Wasser im Bereich von 0°C; und

Fig. 10

eine Darstellung der vom Verdampferaustrittsdruck während des Abtauprozesses durchlaufenen Bereiche.

In the following the invention is illustrated by means of an embodiment and with reference to the accompanying drawings, in which:

Fig. 1

a schematic representation of a compression refrigeration system according to a first embodiment;

Fig. 2

a representation of the flowchart of the method according to the invention;

Fig. 3

a further illustration of the flowchart of the method according to the invention;

Fig. 4

a schematic representation of a refrigerator according to a second embodiment;

Fig. 5

a representation of the flowchart for calculating an average taut temperature;

Fig. 6

a representation of the flowchart as a function of the averaged taut temperature and time;

Fig. 7

a representation of the calculated values for the amplitude as a function of time;

Fig. 8

a representation of the outdoor temperature compensation;

Fig. 9

a representation of the temperature behavior of water in the range of 0 ° C; and

Fig. 10

a representation of the traversed by the evaporator outlet pressure during the Abtaugrozesses areas.

A block diagram of a Kompressionskälteantage is shown in Fig. 1 according to a first embodiment. A refrigeration system consists of the components evaporator 11, compressor 12, condenser 13 and throttle body 15, which are connected by a conduit system through which the refrigerant is passed.

In the embodiment shown in FIGS. 1 to 3, an expansion valve 15 is used as the throttle body 15. Alternatively, a reciprocating engine or a turbine may be used as the throttle body.

By supplying heat to a low temperature level, a medium with a low boiling point ("refrigerant", today mostly ozone harmless CFCs or natural substances) is vaporized in the evaporator 11, the gaseous phase is then compressed in a compressor 12 and thereby heated. Under high pressure, the working fluid releases its heat for use at the condenser 13 (heating water, air flow) and condenses. By an expansion valve 15, the working fluid enters the Tell cycle at low pressure again and is in turn fed to the evaporator 11, at whose output the evaporator pressure is determined by the measuring unit 16.

The temperature difference between the heat source and the refrigerant allows a heat flow to the evaporator 11. Subsequently, the refrigerant vapor is sucked by the compressor 12 and compressed. The temperature of the refrigerant is "pumped" through the temperature level of the heat distribution. At the condenser 13 is again a temperature difference, and there is a heat flow, for heat distribution. The high-pressure refrigerant cools again, condenses and is expanded via an expansion valve 15. The entire process takes place again and is thus in a circular process.

According to the invention, the chiller additionally has a determination unit 21 for determining a first control value W ₁ for the expansion valve 15 as a function of the deviation of an overheating of the refrigerant from a desired overheating. Further, a unit 14 for determining the condenser pressure and a measuring unit 16 for measuring the evaporator pressure is provided. A model 17 for forming a model which compares the mass flow rate at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, a computation unit 18 for calculating a second output value W ₂ for the expansion valve 15 from the evaporator pressure, condenser and refrigerant circuit specific variables, a determination unit 19 for determining a third control value W ₃ for the expansion valve 15 by linking the first control value W ₁ to the second control value W ₂ and a Actuator 20 for setting the expansion valve 15 to the third control value W ₃ is also provided.

During the process of controlling a compression refrigeration system, unit 14 determines condenser pressure and measuring unit 16 measures evaporator pressure at the evaporator outlet. From the evaporation pressure, the evaporation temperature is determined. The formula for calculation is a formula approximation to the dependencies found by measurements on the particular refrigerant used.

From the evaporation temperature and the evaporator outlet temperature, the instantaneous actual overheating of the refrigerant can be derived: From the comparison of the actual overheating with the target overheating, a first control value W ₁ for the expansion valve 15 is determined by means of a controller, to which the opening angle of the expansion valve 15 set and thus the Kalternittellauf is regulated in the circulation. If the actual overheating is greater than the setpoint overheating, the actuator should start up, ie the first actuating signal will increase. If the actual overheating is less than the setpoint overheating, then the actuator should close, that is, the first control signal is smaller. The controller can be designed as a P, PI, I or PID controller.

During the method according to the invention, a second control value W ₂ and third control value W _{3 are} determined in addition to the first control value W ₁ . For this purpose, a model is formed in the unit 17, which compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet. In the arithmetic unit 18, a second control value W ₂ for the expansion valve 15 is calculated on the basis of the model from the evaporator pressure, the Vefflüssigerdruck and cold circle specific variables

The determination unit 19 combines the first control value W ₁ with the second control value W ₂ and in this way determines a third control value W ₃ , to the value of which the expansion valve 15 is adjusted by means of the control unit 20.

FIG. 2 shows how a control circuit for the overheating of the evaporator can be operated, taking into account the predicted control signal magnitude.

In block B1, the pretreatment and evaluation of the sensor signals from the refrigeration circuit. The sensor signals are freed of interference signals (for example 50 Hz hum) by means of low-pass, the sensor time constants are compensated. Furthermore, the actual overheating is calculated from the evaporator outlet temperature and the evaporator pressure and the condenser pressure is calculated from the condenser temperature.

The input signals of block B1 are the evaporator pressure p ₀ , the compressor inlet temperature t _v1 , the evaporator outlet temperature t ₀₂ and the condenser outlet temperature t _c2 .

If no recuperator is installed in the refrigerant circuit, both temperatures (compressor _inlet temperature t _v1 and evaporator _outlet temperature t ₀₂ ) are the same because the evaporator outlet is connected directly to the compressor inlet. If a recuperator is interposed, it increases the refrigerant temperature during heat transfer, and the overheating can be regulated either before or after the recuperator, depending on the design of the refrigeration circuit control.

In block B3, the precalculation of the second actuating signal for the expansion valve is then carried out with the aid of the process values from block B1 using the refrigeration model.

In block B2, a pendulum detection of the signal is performed, and together with block B5 is evaluated by means of the process values from block B1, the operating point of the refrigeration circuit and set a corresponding desired overheating.

In block B4, a controller, the control deviation of overheating (difference between actual overheating ΔT _actual and set overheating ΔT _setpoint ) is fed in and in Output signal influenced by the control deviation is output. In this method step, the first control value is calculated.

Subsequently, the second actuating signal is linked with the aid of the refrigeration-technical model with the control signal influenced by the first control signal to a total control signal. This is advantageously done by multiplication. In this case, the factor formed by the controller output = 1, insofar as there is no deviation.

If a control deviation of the overheating results, the factor formed by the controller output is not equal to 1, and the precalculated actuating signal is corrected accordingly with the aid of the refrigeration technology model. However, other mathematical operations such as addition or weighting are possible.

The precalculated actuating signal passes through block B6 for further processing. Here, the third control signal is, for example, adapted to the control range limits of the expansion valve, and there is also a limitation of the control signal rise in order not to "overtax" the time constant of the refrigerant circuit. By this is meant that it is not necessary control technology when the actuating speed of the actuator exceeds the time constant of the refrigeration circuit by a multiple (for example by a factor of 100). In this case, a very short-term fluctuating control signal would be calculated for very short-term interference (EMC, measurement signal fluctuations, etc.), which would be completely damped by the time constant of the refrigerant circuit, but charged the actuator.

This is particularly advantageous because the valve can not be adjusted infinitely fast and the cooling process does not react arbitrarily fast. Furthermore, block 6 limits the control signal to the physical control range of the valve.

In block B7 is selected depending on the operating state, which signal is forwarded as a control signal to the actuator. In control mode, the mathematically linked and limited control signal is forwarded, such as already stated. Further operating modes are the pump-down operation, a faulting fault or the defrosting operation.

At a specified time after compressor start, it may be necessary that only the precalculated control signal is forwarded to the actuator, because the controller can not provide a meaningful controller output signal due to the highly dynamic processes in the refrigerant circuit.

In special modes such as defrost or standby advantageously a fixed value is forwarded to the actuator.

Block 8 is an evaluation unit, with the aid of which the first actuating signal is evaluated. In the case of a refrigerant shortage detection, it is evaluated whether the first actuating signal in the operating mode control mode exceeds a parameterized value (here a value >> 1) for a minimum period of time. In this case, a lack of refrigerant is detected, brought to the display and possibly causes in block 7, a modified processing of the third actuating signal, for example, emergency operation.

M is the servomotor of the expansion valve, which is coupled with this.

In Fig. 3, the flow chart of the method according to the invention is shown schematically. The process variables used in the calculations are the evaporator pressure p ₀ , the condenser pressure p _c and the associated temperature variables.

By way of example, simplified dependencies for the precalculation of the actuating signal for an expansion valve of a refrigeration circuit of a compression refrigeration machine are described below.

The model is based on the physical background that the refrigerant mass flow at the evaporator inlet (from the expansion valve to the evaporator inlet) in a steady state refrigeration cycle at constant ambient conditions Evaporator) is equal to the refrigerant mass flow at the evaporator outlet (from the evaporator to the compressor)

To model this, the two refrigerant mass flows are equated with their respective influencing variables, which are measured in the refrigerant circuit. Furthermore, physical dependencies in the compressor and expansion valve are included in the modeling.

The mass flow at the evaporator outlet depends on the delivery behavior of the compressor. This is largely determined by the refrigerant pressures on the high-pressure and low-pressure sides of the refrigeration cycle and the degree of delivery influenced by them. The factor const ₁ parameterizes the design-related delivery rate for the refrigerant used in the compressor. This refers to a characteristic operating point, for other operating points deviations are tolerated, which are usually taken from a compressor data sheet or to be determined by laboratory measurements.

wobei $${\mathrm{Liefergrad}}_{\mathrm{Verdichter}}\mathrm{=}(\mathrm{0}\mathrm{,}\mathrm{95}\mathrm{-}\mathrm{\left(}\frac{{\mathrm{p}}_{\mathrm{c}}}{{\mathrm{p}}_{\mathrm{0}}}\mathrm{\right)}\mathrm{*}\mathrm{0}\mathrm{,}\mathrm{1}\mathrm{)}$$

gilt.As a formula for calculating the intake mass flow of the compressor from the evaporator pressure p ₀ and the condenser pressure p _c , including the delivery degree curve (fictitious linearized delivery-grade curve), the following applies: $${\mathrm{Ansaugmasse}}_{\mathrm{compressor}}={\mathrm{<figure-callout\; id="0"\; label="p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{0}\mathrm{*}}\left(\mathrm{0}\mathrm{.}\mathrm{95}\mathrm{-}\left(\frac{{\mathrm{<figure-callout\; id="0"\; label="p\; c\; p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p\; </figure-callout>}}_{\mathrm{c}}}{{\mathrm{p}}_{}}\right)\mathrm{*}\mathrm{0}\mathrm{.}\mathrm{1}\right)*{\mathrm{<figure-callout\; id="1"\; label="const"\; filenames="imgf0005.png"\; state="\{\{state\}\}">const</figure-callout>}}_1$$

in which $${\mathrm{volumetric\; efficiency}}_{\mathrm{compressor}}\mathrm{=}(\mathrm{0}\mathrm{.}\mathrm{95}\mathrm{-}\mathrm{\left(}\frac{{\mathrm{p}}_{\mathrm{c}}}{{\mathrm{p}}_{\mathrm{0}}}\mathrm{\right)}\mathrm{*}\mathrm{0}\mathrm{.}\mathrm{1}\mathrm{)}$$

applies.

The mass flow at the evaporator inlet depends on the mass flow rate at the expansion valve. This is determined by the refrigerant pressures on the high pressure and low pressure side as well as by the central opening cross section of the expansion valve certainly. The opening cross-section is controlled by a control or regulation in the case of electronic expansion valves. In factor const ₂ , the mass flow rate of the expansion valve is parameterized for the refrigerant used. This refers to a characteristic operating point, deviations are tolerated for other operating points.

The formula for calculating the mass flow at the nozzle from the evaporator pressure p ₀ , the condenser pressure p _c and the nozzle cross-section of the expansion valve is as follows: $${\mathrm{mass\; flow}}_{\mathrm{jet}}\mathrm{=}\sqrt{\left({\mathrm{p}}_{\mathrm{c}}\mathrm{-}{\mathrm{<figure-callout\; id="0"\; label="p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{0}}\right)}\mathrm{*}{\mathrm{cross-section}}_{\mathrm{jet}}\mathrm{*}{\mathrm{const}}_{\mathrm{2}}$$

In a refrigeration cycle, in the steady state at constant ambient conditions, the refrigerant mass flow at the evaporator inlet is equal to the refrigerant mass flow at the evaporator outlet. It follows $${\mathrm{mass\; flow}}_{\mathrm{jet}}={\mathrm{Ansaugmasse}}_{\mathrm{compressor}}$$

Equation of the formulas for the mass flows and resolution according to the nozzle cross-section results as a manipulated variable $${\mathrm{Querchnitt}}_{\mathrm{jet}}=\sqrt{\frac{{\mathrm{p}}_{\mathrm{0}}}{\left({\mathrm{p}}_{\mathrm{c}}\mathrm{-}{\mathrm{<figure-callout\; id="0"\; label="p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{0}}\right)}}\mathrm{*}{\mathrm{volumetric\; efficiency}}_{\mathrm{compressor}}*\mathrm{const}$$

The relationship between nozzle area and control signal for an expansion valve with a conical nozzle needle consists of: $${\mathrm{manipulated\; variable}}_{{\mathrm{expansion\; valve}}_{-}\mathrm{rel}}=1-\sqrt{\left(1-{\mathrm{cross-section}}_{{\mathrm{jet}}_{-}\mathrm{rel}}\right)}$$

In the following it is described how, in dependence on an exemplary valve characteristic curve with offset, the nozzle cross-section can be replaced by a setting step. $${\mathrm{Sch}\begin{array}{}\end{array}r\begin{array}{}\end{array}\mathrm{itt}}_{\mathrm{Expansionsv}\mathrm{,}}=1-\sqrt{\left(1-\frac{{\mathrm{p}}_{\mathrm{0}}}{\sqrt{\left({\mathrm{p}}_{\mathrm{c}}\mathrm{-}{\mathrm{<figure-callout\; id="0"\; label="p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{0}}\right)}}*{\mathrm{volumetric\; efficiency}}_{\mathrm{Verd}\mathrm{,}}*\mathrm{const}\right)}+{\mathrm{offset}}_{\mathrm{Expansionsv}\mathrm{,}}$$

The factor of the degree of compressor delivery and the relationship between the nozzle cross-section and control signal for an expansion valve can be approximately integrated into Exp _{valve characteristics} and in const: $${\mathrm{Sch}\begin{array}{}\end{array}r\begin{array}{}\end{array}\mathrm{itt}}_{\mathrm{expansion\; valve}}={\left(\frac{{\mathrm{p}}_{\mathrm{0}}}{\sqrt{\left({\mathrm{p}}_{\mathrm{c}}\mathrm{-}{\mathrm{<figure-callout\; id="0"\; label="p"\; filenames="imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{0}}\right)}}\right)}^{\mathrm{Exp}}*\mathrm{const}+{\mathrm{offset}}_{\mathrm{Valve}}$$

The evaporator pressure and the condenser pressure are measured as process variables in the refrigerant circuit. In a further embodiment according to the invention, the condenser pressure can be calculated from the condenser temperature by means of refrigerant data.

As fixed variables enter into the model: the exponent Exp, the offset and the cold-circle-specific constant const, these fixed variables are dependent on the respective components of a refrigeration circuit. As a fixed size, the offset of the expansion valve, which describes the number of steps until the first opening. The exponent maps both the function of the nozzle cross section over the output level and the function of the delivery rate of the compressor. The exponential function formed by the exponent approximates the refrigeration cycle component-specific functions

The parameterization of the model is carried out via a single constant of the constant of a cold-circuit. This parameter forms the sum of the parameters in the compressor, condenser, expansion valve and evaporator, which is determined by laboratory measurements or calculation. As a further advantageous embodiment, the cold-zone-specific constant const during operation of the refrigeration circuit be adapted so that the calculation of the expansion valve steps due to the refrigeration cycle model is becoming more accurate.

A further advantageous embodiment of the method is to adapt the cold-cycle-specific constant const ascertained for example in laboratory tests in the course of operation in such a way that the control signal obtained with the aid of the refrigeration-technical model including the constant const adapts optimally to the refrigeration process. In this case, the corrections of the controller required by a control deviation in block B4 are minimal, the control is very accurate.

Furthermore, a conclusion can be drawn from the behavior of the closed loop on the present lack of refrigerant. The relationships described in the refrigeration model are based on the assumption that there is sufficient refrigerant to operate the refrigeration circuit. If refrigerant escapes, e.g. due to leaks or if the refrigerant circuit is insufficiently filled before commissioning or after a component change, a control value of the expansion valve deviating from the refrigeration model is required to set the overheating at certain operating points.

This manifests itself in operation in that the control signal (block B3) predefined by the cooling-technical model has to be corrected to a greater extent by the controller (block B4). This in turn means that setting the desired overheating requires a much larger actuating signal than predicted, ie. with multiplicative connection of the control signals, the controller output signal is substantially greater than 1 when the control loop is steady.

Particularly advantageous is an embodiment of the method according to the invention, in which a lack of refrigerant detected and appropriate measures are triggered when in the steady state of the control loop in the control mode, the control signal, which is the output of the superheat controller is detected for a set time above a predetermined value.

The expansion valve in a preferred embodiment can be adapted to each of the three control values, depending on the operating mode, in order to optimally adapt the mode of operation to the respective operation.

Although a compression refrigeration system has been described in accordance with FIGS. 1 to 3, the principle of the first embodiment can also be applied to an absorption refrigeration machine.

The operation of a compression refrigeration system according to the second embodiment is shown schematically in FIG. 4. The refrigeration system according to the second embodiment can be combined with a refrigeration system according to the first embodiment or it can be operated as a standalone system. A refrigeration system comprises an evaporator 111, a compressor 112, a condenser 113 and a throttle body 115, which are connected by a conduit system through which the coolant is passed. By supplying heat at a low temperature level, a medium with a low boiling point ("refrigerant", today mostly ozone-harmless CFCs or natural substances) is vaporized in the evaporator 111, the gaseous phase then compressed in a compressor 112 and thereby heated. Under high pressure, the working fluid releases its heat for use at the condenser 113 (heating water, air flow) and condenses. By a throttle body (expansion valve 115), the working fluid enters the partial circuit at low pressure again and in turn is fed to the evaporator 111, at whose output the evaporator outlet pressure is determined by the measuring device 16.

The temperature difference between the heat source and the refrigerant allows a heat flow to the evaporator 111, Subsequently, the refrigerant vapor is sucked by the compressor 112 and compressed. The temperature of the refrigerant is "pumped" through the temperature level of the heat distribution. At the condenser 113 is again a temperature difference and there is a heat flow, for heat distribution. The high-pressure refrigerant cools again, condenses and is via an expansion valve 115 relaxed. The entire process takes place again and is thus in a circular process.

The refrigerating machine according to the invention further comprises a measuring unit 116 for measuring the evaporator outlet pressure, a determining unit 117 for calculating a melting temperature from the evaporator outlet pressure, a first determining unit 118 for determining a first difference from the melting temperature and a melting temperature reference value, a second determining unit 119 for determining a second difference from the evaporator outlet pressure and a cut-off pressure, and a defrost unit 120 for initiating a defrost operation if the first difference exceeds a temperature threshold and terminating defrost if the second difference is below a pressure threshold.

The measuring device 116 detects at the outlet of the evaporator 111, the pressure of the refrigerant, which is passed from the evaporator to the compressor. From this measured evaporator outlet pressure, the tau temperature is calculated in the determination unit 117, which in turn flows into further arithmetic operations. The arithmetic unit can perform further computation steps, as explained below in the flowchart in FIG.

In the determination unit 118, a first difference is determined from the tare temperature calculated in the determination unit 117 and a taut temperature reference value. If this first difference exceeds a temperature limit, the defrost unit 120 initiates the defrost process for the evaporator. The flowchart of Fig. 6 illustrates this process step in detail.

In the determination unit 119, a second difference is formed from the evaporator outlet pressure and a shut-off pressure. The defrost unit 120 stops the defrosting process if the second difference falls below a pressure threshold. This process step will be explained with reference to FIG. 10.

In FIG. 5, algorithmic calculation steps for determining an averaged taut temperature of the method according to the invention are shown schematically by way of example. Since the evaporator temperature signal can oscillate depending on the operating condition, it must be averaged and filtered. In order to obtain the most stable signal, further signals are generated from the evaporation temperature.

• die aktuelle Verdampfertemperatur θ₀, die aus dem gemessenen Verdampferausgangsdruck berechnet wird,
• die maximale Amplitudentemperatur θ_{max_amp,n},
• die minimale Amplitudentemperatur θ_{min_amp,n}, und
• die Mitteltemperatur θ_mittel,n.

In order to perform the calculation steps, the following values are required as input variables:

• the actual evaporator temperature θ ₀ , which is calculated from the measured evaporator outlet pressure,
• the maximum amplitude _temperature θ _{max_amp, n} ,
• the minimum amplitude _temperature θ _{min_amp, n} , and
• the mean temperature θ _{middle, n} .

By means of the calculation steps, the temperatures θ _{i, n + 1 of} the following cycle are calculated from the variables θ _{i, n of} the current cycle, wherein the mean temperature θ _{middle, π + 1 is} the essential quantity to detect a defrost requirement. Furthermore, C _Ab as a factor for the decay of the signal and C _on as a factor for the sounding of the signal into the calculation.

The factors C _Ab and C _Auf are determined as follows. In general, the factors are to be dimensioned so that in the case of a cyclical fluctuation of the overheating, ie oscillation, the minimum and the maximum amplitude temperature the periodic maxima and the periodic minima of overheating depict. If the oscillation stops, the minimum and maximum amplitude temperature should be adapted to the periodically decaying minima and maxima of overheating.

The factor C _Auf is to be dimensioned so that the minimum and maximum amplitude temperature can follow the gradient of fluctuating overheating so that the maxima and minima can be followed almost unattenuated.

The time constant for the fade should be a fraction of the oscillation time constant of commuting, for example a quarter of this.

The factor C _Ab is to be dimensioned so that the minimum and the maximum amplitude temperature between two maxima or minima is largely retained and does not completely decay, so that an envelope is described with both amplitude temperatures. The time constant for the fade should be a multiple of the swing time constant of the pendulum, for example, twice this.

The oscillation time constant of oscillation is approx. 2 to 10 minutes, depending on the cooling circuit. If the time constants are dimensioned for fading up and down, the factors can be calculated after the controller iterative time.

The maximum amplitude of the temperature θ _{max_amp, n1 is} the medium temperature θ _{medium, n} going _from C and as factors for the decay of the signal in the calculation of a temperature T ₁ a: $${\mathrm{<figure-callout\; id="1"\; label="T"\; filenames="imgf0005.png"\; state="\{\{state\}\}">T</figure-callout>}}_{\mathrm{1}}\mathrm{=}\left(\mathrm{1}\mathrm{-}{\mathrm{C}}_{\mathrm{From}}\right)\mathrm{\cdot }{\mathrm{A}}_{{\mathrm{Max}}_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}\mathrm{+}{\mathrm{C}}_{\mathrm{From}}\mathrm{\cdot }{\mathrm{\theta }}_{\mathrm{by\; means\; of}\mathrm{.}\mathrm{n}}$$

The actual evaporator temperature θ ₀ , the maximum amplitude _temperature θ _{max_amp, n} and C _on go into the calculation of a temperature T ₃ as factors for the sounding of the signal: $${\mathrm{<figure-callout\; id="3"\; label="T"\; filenames="imgf0003.png,imgf0005.png"\; state="\{\{state\}\}">T</figure-callout>}}_3\mathrm{=}\left(\mathrm{1}\mathrm{-}{\mathrm{C}}_{\mathrm{On}}\right)\mathrm{\cdot }{\mathrm{A}}_{{\mathrm{Max}}_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}\mathrm{+}{\mathrm{C}}_{\mathrm{On}}\mathrm{\cdot }{\mathrm{<figure-callout\; id="0"\; label="\theta "\; filenames="imgf0005.png"\; state="\{\{state\}\}">\theta </figure-callout>}}_0$$

The maximum amplitude temperature for a decaying and an evanescent signal is determined. If the actual evaporation temperature is greater than the maximum amplitude temperature, it approaches the instantaneous value of Evaporation temperature with the factor C _up . at the same time it approaches the mean temperature with the factor C _Ab : $${\mathrm{\theta }}_{{\mathrm{Max}}_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}\mathrm{+}\mathrm{1}}\mathrm{=}{\mathrm{<figure-callout\; id="1"\; label="T"\; filenames="imgf0005.png"\; state="\{\{state\}\}">T</figure-callout>}}_{\mathrm{1}}\mathrm{+}{\mathrm{<figure-callout\; id="0"\; label="WENN\theta "\; filenames="imgf0005.png"\; state="\{\{state\}\}">WENN\theta </figure-callout>}}_{\mathrm{0}}\mathrm{>}{\mathrm{\theta }}_{{\mathrm{Max}}_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}{\mathrm{DANNT}}_{\mathrm{3}}$$

Analog computing steps are performed to determine the minimum amplitude temperature. The minimum amplitude temperature θ _{min_amp.n,} the mean temperature θ _{medium, n} and C _From go as factors for the decay of the signal in the calculation of a temperature T ₂ a: $${\mathrm{T}}_{\mathrm{2}}\mathrm{=}\mathrm{(}\mathrm{1}\mathrm{-}{\mathrm{C}}_{\mathrm{From}}\mathrm{)}\mathrm{\cdot }{\mathrm{\theta }}_{{\min }_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}\mathrm{+}{\mathrm{C}}_{\mathrm{From}}\mathrm{\cdot }{\mathrm{\theta }}_{\mathrm{medium}\mathrm{.}\mathrm{n}}$$

The actual evaporator temperature θ ₀ , the minimum amplitude _temperature θ _{min_amp.n} and C _on go as factors for the sounding of the signal in the calculation of a temperature T ₄ : $${\mathrm{<figure-callout\; id="4"\; label="T"\; filenames="imgf0005.png,imgf0008.png"\; state="\{\{state\}\}">T</figure-callout>}}_4\mathrm{=}\mathrm{(}\mathrm{1}\mathrm{-}{\mathrm{C}}_{\mathrm{On}}\mathrm{)}\mathrm{\cdot }{\mathrm{\theta }}_{{\min }_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}\mathrm{+}{\mathrm{C}}_{\mathrm{On}}\mathrm{\cdot }{\mathrm{<figure-callout\; id="0"\; label="\theta "\; filenames="imgf0005.png"\; state="\{\{state\}\}">\theta </figure-callout>}}_0$$

Finally, the minimum amplitude temperature is determined. If the actual evaporation temperature is less than the minimum amplitude temperature, it approximates the instantaneous value of the evaporation temperature with the factor C _up , at the same time it approaches the mean temperature with the factor C _Ab : $${\mathrm{\theta }}_{{\min }_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}\mathrm{+}\mathrm{1}}\mathrm{=}{\mathrm{T}}_{\mathrm{2}}\mathrm{+}{\mathrm{<figure-callout\; id="0"\; label="WENN\theta "\; filenames="imgf0005.png"\; state="\{\{state\}\}">WENN\theta </figure-callout>}}_{\mathrm{0}}\mathrm{<}{\mathrm{\theta }}_{{\min }_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}}{\mathrm{DANNT}}_{\mathrm{4}}$$

From the calculated values for the minimum and the maximum amplitude temperature, the mean temperature T _{5 is} calculated. The mean value is calculated from the average of maximum and minimum amplitude temperature. $${\mathrm{<figure-callout\; id="5"\; label="T"\; filenames="imgf0008.png"\; state="\{\{state\}\}">T</figure-callout>}}_{\mathrm{5}}\mathrm{=}\mathrm{0}\mathrm{.}\mathrm{5}\mathrm{\cdot }\left({\mathrm{\theta }}_{{\min }_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}\mathrm{+}\mathrm{1}}\mathrm{+}{\mathrm{\theta }}_{{\mathrm{Max}}_{\mathrm{-}}\mathrm{amp}\mathrm{.}\mathrm{n}\mathrm{+}\mathrm{1}}\right)$$

With the calculated mean temperature and additional reference variables, the defrosting requirement of a chiller can be identified. In this case, T _{5 is} the mean temperature θ _medium .

In Fig. 6 by way of example the A _{flow of} the inventive method for Abtaubedarferkennung is shown schematically. The method according to the invention begins with the sequence of operations after an initial blocking time t _blocking which is between 5 and 30 minutes, preferably between 10 and 15 minutes, which corresponds to the time frame in which the system settles.

In the method according to the invention, the evaporator outlet pressure is first measured and the tau temperature is calculated therefrom. The taut temperature is then averaged according to the flowchart in FIG.

A defrost requirement is recognized in the method according to the invention in that a first difference between an averaged taut temperature and a taut temperature reference value exceeds a temperature limit value. The tew temperature reference value can be determined from the mean temperature.

wobei I der erste Messzyklus nach Ablauf der Sperrzeit t_Sperr ist.The value of the temperature reference value θ _{ref_max} preferably corresponds to the maximum value of the mean temperatures of the current heating cycle. At the beginning of the heating cycle, the maximum value formation is inactive, since the cooling circuit has not yet settled and the amplitudes are disproportionately large: $${\mathrm{\theta }}_{{\mathrm{ref}}_{\mathrm{-}}\mathrm{Max}\mathrm{.}\mathrm{n}\mathrm{+}\mathrm{1}}\mathrm{=}{\mathrm{Max}}_{\mathrm{i}\mathrm{=}\mathrm{1}}^{\mathrm{n}}\left({\mathrm{\theta }}_{\mathrm{medium}\mathrm{.}\mathrm{i}}\right)$$

where I is the first measurement cycle after expiration of the blocking time t _lock .

As input variables, the time t and the average taut temperature θ _medium are included in the method illustrated in FIG. In method step S31, the following step takes place $$\mathrm{IF}\left(\mathrm{T}\mathrm{>}{\mathrm{T}}_{\mathrm{lock}}\right)\mathrm{AND}\left({\mathrm{\theta }}_{\mathrm{medium}}\mathrm{>}{\mathrm{\theta }}_{{\mathrm{ref}}_{\mathrm{-}}\mathrm{Max}}\right)\mathrm{THEN}\left({\mathrm{\theta }}_{{\mathrm{ref}}_{\mathrm{-}}\mathrm{Max}}\mathrm{=}{\mathrm{\theta }}_{\mathrm{medium}}\right)$$

In this method step, therefore, after the initial blocking time, the maximum averaged taut temperature up to the time t is set as the taut temperature reference value.

The _taut temperature _{reference value} θ _{ref_max} determined from the method step S31 and the averaged _taut temperature θ _middle enter the method step S32 in which the difference of these two variables is calculated and compared with a taut temperature _limit value T _limit : $$\mathrm{IF}\left({\mathrm{\theta }}_{{\mathrm{ref}}_{\mathrm{-}}\mathrm{Max}}\mathrm{-}{\mathrm{\theta }}_{\mathrm{medium}}\mathrm{>}{\mathrm{T}}_{\mathrm{border}}\right)\mathrm{THEN}\left[\mathrm{S}\mathrm{33}\right]$$

If the difference is thus greater than a temperature limit, method step 33 is initiated. This method step contains a loop which checks how often the difference from step S32 has exceeded the temperature _limit T limit. The defrost process in step S34 is initiated if step S33 is initiated at least once. in a preferred embodiment, five times having received positive information from step S32

FIG. 7 shows the course of the calculated values for the amplitude as a function of time. On the x-axis the time course is shown, on the y-axis the Amplitudes. It can clearly be seen from the illustration that the tau temperature θ ₀ calculated from the measured evaporator pressure oscillates and is enclosed by the two calculated values for the minimum and the maximum amplitude temperature. The mean value θ _mean lies between the minimum and maximum amplitude temperature. Above these temperature _values is the temperature _{reference value} θ _{ref_max} . which only changes if the mean temperature exceeds the temperature reference value

Fig. 8 shows the dependencies in the outdoor temperature compensation of the tau temperature θ ₀ . In the method of outdoor temperature compensation, the outside temperature θ is measured _outside and included in the calculation of the tau temperature. A tare temperature θ _corrected by the outside temperature θ _outside is calculated by forming the difference of a second tau temperature θ ₀ and the outside temperature θ _outside . Thus, the temporal behavior of the second taut temperature θ ₀ relative to the outside temperature θ _outside rated. The corrected melting temperature θ _corrected enters the method according to the invention as the first melting temperature.

Fig. 9 is a graph showing the temperature behavior of water in the range of 0 ° C. On the y-axis is the temperature, on the x-axis the time course is shown. Water has the special property that during the thawing process it remains at the 0 ° C temperature level for a longer time until the temperature finally rises further. This effect is exploited to determine the optimal point at which the entire ice has melted. Only when all the ice has melted on the evaporator surface, the temperature rises from the temperature plateau of 0 ° C, which manifests itself in an increase in the evaporator pressure. Thus, the defrost process is terminated only when the evaporator is free of ice

FIG. 10 shows the various pressure ranges which the evaporator outlet pressure p ₀ undergoes during the defrosting process. For this purpose, the temperature is plotted on the y-axis and the time on the x-axis.

In zone I, the temperature rises to "melt pressure", II denotes the temperature plateau to "melt pressure" and III shows the temperature increase after defrost, the pressure reaches the cut-off, which is about 2 bar above the evaporator outlet pressure corresponding to the melting temperature of the ice, the defrost will stop.

According to one aspect of the second embodiment of the present invention, there is provided a method of controlling a defrosting operation of an evaporator of a refrigerating machine (such as a refrigerating machine according to the first embodiment). Alternatively or additionally, the method according to the second embodiment can be operated independently of the refrigerating machine according to the first embodiment. For this purpose, an evaporator outlet pressure is measured. A teat temperature is determined based on the evaporator outlet pressure. A first difference from the first teat temperature and a tau temperature reference value is determined. A defrost operation is initiated if the first difference exceeds a temperature limit. Defrosting is accomplished by determining a second difference from the evaporator outlet pressure and a shutoff pressure and terminating defrost if the second difference is less than a pressure threshold.

The evaporator outlet pressure shows a characteristic of the refrigeration cycle over time, from which can be concluded on a non-performance optimized operation. The dew temperature can be calculated from the evaporator outlet pressure. If the surface structure of the evaporator ices up, the efficiency deteriorates, the evaporator pressure drops and thus also the calculated peat temperature. If the difference between the calculated peat temperature and a reference temperature exceeds a temperature limit, a defrost requirement can be detected and a defrost process initiated. During defrost, the evaporator outlet pressure is still monitored and compared to a shutdown pressure. If the difference between these two pressures becomes sufficiently small, the evaporator surface is sufficiently defrosted and the defrosting operation is terminated.

The advantages of this method are that a significantly increased detection reliability is given. The calculation is preferably carried out regularly during the operation of the chiller, which also temperature jumps of the outside air can be perceived quickly enough and is responded accordingly. When using pressure sensors, a minimum inertia in the detection process is ensured under a low defrost energy requirement.

The components in the circuit of the chiller can be reused without having to be exchanged for expensive and sensitive electronic units.

Particularly preferred in the inventive method according to the second embodiment is an embodiment in which the first taut temperature is calculated from a difference between a second tau temperature and an outside temperature, wherein the second tau temperature is calculated from the evaporator outlet pressure and the outside temperature is measured. This evaluates the temporal behavior of the peat temperature relative to the outside temperature

In a preferred embodiment, the temperature limit, cut-off pressure and pressure limit are fixedly defined for a particular installation. However, they may also be e.g. be adaptable to special external conditions, in particular via manual adjustment or an adaptation device.

Preferably, the tew temperature reference value is determined by the maximum averaged taut temperature to ensure an adaptive method.

In order to determine the optimum point for terminating the defrosting operation, during the defrosting operation, the evaporator outlet pressure is compared with a shutdown pressure. The switch-off pressure is around 1 to 3 bar, preferably around 2 bar, above the evaporator outlet pressure corresponding to the melting temperature of the ice.

In the defrost detection method according to the second exemplary embodiment, only one measurement of the evaporator outlet pressure is required, from which the remaining data required for the evaluation are determined.

Claims (16)

Method for controlling a compression refrigerating machine with a refrigerant, an evaporator (11), a pressure increasing unit (12), a condenser (13) and a throttling device (15), comprising the steps of: a) determining a first control value for the throttle element (15) as a function of the deviation of an overheating of the refrigerant from a desired superheating, b) determining the condenser pressure, c) measuring the evaporator pressure, d) forming a model which compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, e) calculating a second control value for the throttle body (15) on the basis of the model from the evaporator pressure, the condenser pressure and Kättekreisspezifischen sizes, f) determining a third control value for the throttle element (15) by linking the first control value with the second control value and g) setting the throttle body (15) to the third control value. Method according to claim 1,
characterized in that step a) comprises the following substeps: - Measuring the evaporator pressure and the refrigerant temperature at the evaporator outlet Calculating the evaporation temperature from the evaporator pressure and refrigerant specific data, Determining an actual overheating of the refrigerant at the evaporator outlet from the difference between the refrigerant temperature and the evaporating temperature, Determining the deviation of the actual overheating from a desired overheating, and - Determining a first control value for the throttle body as a function of the deviation of the actual overheating of the desired superheat. Method according to one of the preceding claims,
characterized in that enter into the model of the evaporator pressure and the condenser pressure as process variables and an offset of the throttle body, a kaltekreisspezifische constant and an exponent as cold circle specific variables. Method according to one of the preceding claims,
characterized in that instead of the process variable of the evaporator pressure, the process variable enters the evaporator inlet temperature in the model, wherein it is converted by means of the refrigerant characteristic into the evaporator pressure. Method according to one of the preceding claims,
characterized in that the condenser temperature is measured and the condenser pressure is calculated from the condenser temperature. Method according to one of the preceding claims,
characterized in that the condenser pressure is measured. Method according to one of the preceding claims, wherein a defrosting of an evaporator (11) of a chiller is carried out with the following steps: a) measuring the evaporator outlet pressure, b) determining a first melting temperature based on the evaporator outlet pressure, c) determining a first difference from the first tau temperature and a tau temperature reference value, d) initiating a defrost operation if the first difference exceeds a temperature limit, the defrosting operation comprising the following steps d1) determining a second difference from the evaporator outlet pressure and a shutdown pressure, d2) Ending the defrosting process, if the second difference falls below a pressure limit value. Method according to claim 7,
characterized in that the first teat temperature is calculated from a difference between a second teat temperature and an outside temperature, wherein the second teat temperature is calculated from the evaporator outlet pressure and the outside temperature is measured. Method according to one of claims 7 to 8,
characterized in that the incoming in the subtraction first taut temperature in step c) is one of at least two calculated in step b) thawing temperatures taut temperature. Method according to one of claims 7 to 9,
characterized in that the steps a) to d2) are carried out regularly, in particular continuously. Method according to one of claims 7 to 10,
characterized in that the temperature limit is fixed. Device with a refrigeration cycle, with a refrigerant, an evaporator (11), a pressure-increasing unit, a condenser (13) and a throttle body (15), with a determination unit for determining a first control value for the throttle element as a function of the deviation of an actual overheating of the refrigerant from a desired overheating, a unit for determining the condenser pressure, a measuring unit for measuring the evaporator pressure, a model unit for forming a model, which compares the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet, an arithmetic unit for calculating a second control value for the throttling device on the basis of the model of the evaporator pressure, the condenser pressure and kättekreisspezifischen sizes, - A determination unit for determining a third control value for the throttle body by linking the first control value with the second control value and - An adjusting unit for adjusting the throttle body to the third control value. The device of claim 12, further comprising: - A measuring unit for measuring the evaporator outlet pressure. a determining unit for determining a first melting temperature based on the evaporator outlet pressure, a first determination unit for determining a first difference between the first tau temperature and a tau temperature reference value, a second determination unit for determining a second difference from the evaporator outlet pressure and a shut-off pressure and a defrost unit for initiating a defrost operation if the first difference exceeds a temperature threshold and terminating the defrost operation if the second difference is less than a threshold pressure. Device according to claim 13,
characterized in that the apparatus comprises a measuring unit for measuring the outside temperature and a calculating unit for calculating the first melting temperature from a difference of a second melting temperature and the outside temperature, wherein the second melting temperature is calculated from the evaporator outlet pressure. Device according to claim 13,
characterized in that the chiller is a heat pump. Device according to claim 13,
characterized in that the chiller is a ventilation system, in particular an air conditioner

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